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2013-01-0340
A New Control Strategy of Wet Dual Clutch Transmission (DCT)
Clutch and Synchronizer for Seamless Gear Preselect
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Affiliation (Do NOT enter this information. It will be pulled from participant tab in MyTechZone)
Copyright © 2012 SAE International
ABSTRACT
In this paper, a new gear preselect strategy of wet dual clutch
transmission (DCT) is proposed by delaying the gear preselect
to the time near the power shift time in order to create a
seamless gear preselect. Software in the loop (SiL) method is
used to investigate the proposed strategy. High fidelity 13
degree of freedom lumped inertia drivetrain models, i.e.,
synchronizer model and electro hydraulic model, are
constructed and validated using laboratory test data. A closed
loop control for electro hydraulic system is developed to
control the clutch and synchronizer considering the optimum
synchronizer trigger time. Furthermore, the model used in this
study is constructed considering the valve spool dynamic in
electro hydraulic system and the sleeve dynamic including
axial drag force in synchronizer system. The genetic algorithm
has been used to obtain the optimum trigger time which is
close to the power shift time to sustain uninterrupted torque
during gear shifting. A comparison between the conventional
sport mode gear preselect strategy with the proposed gear
preselect strategy is performed. The obtained results of the
proposed strategy show a fuel reduction of 0.8% in New
European Driving Cycle (NEDC), within the allowable limits
of the shift time.
1. INTRODUCTION
Recently, many high performance vehicles have been using
dual clutch transmission (DCT) as their transmission to ensure
fast gear shifting with uninterrupted torque for achieving
higher acceleration and speed. On the other hand, the
requirement to reduce fuel consumption is more strictly stated
by the official regulator that typically will diminish the vehicle
performance if the subsequent improvement of the vehicle is
not further made. Thus, the vehicle manufacturers are forced
to increase the efficiencies of all vehicle components to meet
these new regulations. As a complex system, the improvement
of DCT efficiency can be achieved by either improving the
software system as a relatively easy and low-cost method or
changing the DCT hardware component which is considerably
expensive.
Structurally, DCT is a merging of two manual transmissions
which are well known for its robust construction. DCT
employs two separate clutches which hold different gears. The
first clutch holds the odd 1st, 3
rd, 5
th, 7
th gears, whereas the
second clutch grasps the even 2nd
, 4th
and 6th
gears. In order to
work properly, DCT utilizes an intelligence system to control
the actuator mechanism through the electrohydraulic system.
Therefore, the gear in DCT can be shifted automatically.
However, most of DCT still provide the manual gear shift
mode which is preferable to the enthusiast drivers.
Many researchers have been continuously working on the
DCT for better understanding of its system [1-31]. Alvermann
et al. had proposed the clutch control in gear shifting for better
comfort [7], whereas the synchronizer working load due to the
drag torque in transmission was investigated by Walker et al.
[2, 4, 28, 29]. Furthermore, the electrohydraulic system for
gear shift actuator and the shift schedule to increase the
drivability and DCT quality were explored and optimized by
Mustafa et al. and Liu et al., respectively [5-6, 19].
Differentiating from the previous researches, an enhancement
of the wet DCT efficiency is proposed and performed in this
work by improving the gear preselect strategy by knowing the
limitation of DCT mechanical construction and
electrohydraulic system. The obtained results of the proposed
method are also compared with those of the conventional sport
mode gear preselect strategy.
2. GEAR PRESELECT STRATEGY
The DCT gear shift process is merely a simultaneously engine
torque transfer from current closed clutch to the next open
clutch to prevent torque interruption. This mechanism is
similar to that of the automatic transmission which is known
as clutch to clutch power shift, but completely different with
that of the manual transmission i.e., disengaging from one
current gear and then engaging to the next gear. Moreover,
gear preselect can be defined as a gear engagement process of
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one constant mesh gear to the open clutch through the
synchronizer system before the gear shift. This method is used
to guarantee the readiness of an open clutch for receiving the
engine torque from the current closed clutch and
simultaneously transmitting it to the wheel through new gear
ratio. Each clutch is only possible to be engaged with just one
gear at a time. Furthermore, only one of the dual clutches may
be engaged to the engine in order to avoid serious gearbox
damage. In most DCTs, gear preselect is automatically done
before the gear shift by DCT system without any inputs from
the driver. It possibly will cause some dilemmas in manual
gear shift mode when the preselected gear that automatically
done by DCT system is different with the next gear which is
desired by the driver resulting in a longer gear shift time. To
overcome this dilemma, some DCT using so-called manual
gear preselect in manual gear shift mode that manually
selected by the driver before the gear shift to guarantee the
right next gear has been preselected before the gear shift, this
method however will increase the driver task load.
To handle a high engine torque, DCT is using a wet clutch
instead of a dry clutch with the aim of dissipating more heat
generated during clutch engaging process and minimizing the
clutch pad wear to extend the gearbox lifetime. However, wet
clutch DCT also has some drawbacks compared to the dry
clutch DCT, e.g., it needs more energy to pump the clutch
lubrication oil which also produce a viscous drag in case of a
slip between clutch pad set.
Figure 1. Measured rotational speeds of engine, odd clutch,
and even clutch for (a) sport driving mode and (b) comfort
driving mode.
Fig. 1 shows the measured results of the rotational speeds of
engine, odd clutch, and even clutch for two different DCT
driving modes. From this characterization, it is clearly shown
that the gear shift of the sport mode has a higher engine rpm
and a shorter shifting duration than that of the comfort mode.
The different gear shifting durations are intended to be used in
different purposes. The shorter gear shifting duration is well
suited to the sport mode for having a quick response of the
gear shifting. Whereas, the comfort mode is likely to have a
longer shifting time in order to reduce the gear shift jerk.
Moreover, the open clutch will stick and rotate together with
clutch pack before gear preselect taking place in the comfort
driving mode as shown in Fig. 1(b). Nevertheless, after the
gear preselect has been done, the open clutch will rapidly
change its rotation speeds to be either higher or lower than that
of the clutch pack to prepare for the upcoming shift, i.e.,
downshift or upshift. The different gear preselect strategies for
both sport and comfort driving modes are also depicted in Fig.
1. In the sport mode, gear preselect is performed only after the
gear shift has been finished to shorten the next gear shift time.
This mechanism is opposite to that of the comfort mode where
the gear preselect is executed only before the gear shift taking
place. By comparing those two methods, the sport driving
mode obviously exhibits higher drag losses in terms of the
strategy for gear preselect compared to the comfort driving
mode due to the extended slip clutch time.
As an alternative for gear preselect systems, a new seamless
gear preselect is proposed in this study. The working principle
of this gear preselect strategy is similar with that of the
comfort mode. However, instead of postponing the clutch to
clutch power shift after gear preselect is completely done, this
seamless strategy is achieved by almost simultaneously
activating the synchronizer engage action with clutch to clutch
power shift in order to reduce the gap time between gear
preselect and the gear shift as shown in Fig. 2. The action
sequences of the seamless gear preselect method can be
described as the synchronizer engagement to preselect the
wanted next gear, followed by the clutch to clutch power shift
and finally ended by synchronizer disengaging from the
previous gear.
Figure 2. The illustration of the rotational speed of engine,
odd clutch and even clutch for seamless gear preselect
strategy in upshift from even to odd gear.
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The seamless gear preselect strategy is beneficial, particularly
for manual driving mode because the gear preselect is not able
to be further performed automatically by DCT system before
the clutch to clutch power shift. This seamless gear preselect
can only be executed after the driver gear shift command, i.e.,
downshift or upshift, is existed and this will overcome the
dilemma of gear preselect in manual gear shift mode as
mentioned before. Furthermore, the gear shift response time of
the seamless gear preselect will exhibit a slightly longer time
in the manual driving mode compared to the manual sport
mode. Nevertheless, for wet DCT, it will gain the other
advantages, e.g., less clutch drag torque and reduced work
load of DCT system for preparing the next wanted gear by its
preselection.
3. DRIVETRAIN MODEL
The vehicle studied in this paper is BMW M3 E92 model with
7 speed DCT designed by Getrag. The general technical data
of the vehicle engine and transmission are shown in Table 1.
The construction of the DCT comprises four types of shafts
(i.e., counter, hollow, solid, and output shafts), two types of
clutches (i.e., even and odd clutches), and synchronizers as
illustrated in Fig. 3. The solid input shaft connects the first
larger diameter clutch to the set of odd gears while the hollow
input shaft connects the second smaller diameter clutch to the
set of even gears. Two double cone synchronizer systems are
positioned in the counter shaft for the 1st, 2
nd, 3
rd and reverse
gears, while one single cone synchronizer system is positioned
in the solid input shaft for 5th
and 7th
gears and one single cone
synchronizer positioned in hollow input shaft for 4th
and 6th
gears. By using this configuration setup, the engine torque can
flow from one of the two input shafts to the output shaft
through the counter shaft during driving, with the exception
for 7th
gear which is direct torque flow from solid input shaft
to the output shaft with a 1:1 speed ratio.
Figure 3. The structure of DCT Getrag Powershift
Transmission 7DCI600
Table 1. Technical data of BMW M3 Engine and DCT
Getrag Powershift Transmission 7DCI600
BMW M3 E93 Engine Specification Engine type S65B40 naturally aspirated
Capacity/configuration 4.0 liter/V8
Max power 414 bhp @ 8300 rpm Max torque 400 Nm @ 3900 rpm
Weight (dry) 202 kg DCT Getrag 7DCI600 Specification
Maximum input speed 9200 rpm Maximum torque 600 Nm
Weight (dry) 79 kg
Installation length 660 mm Max. vehicle/trailer mass 2500 kg/4500 kg
Gear spread ratio 4.8 Clutch type Wet coaxial normally open
Actuator Electrohydraulic system
Synchronizer System 1
st, 2
nd, 3
rd & reverse gears Double cone synchronizer
4th
, 5th
, 6th
& 7th
gears Single cone synchronizer
3.1. Drivetrain Lumped Inertia Model
The diagram of powertrain model developed for this study is
shown in Fig. 4. This 13 degree of freedom lumped inertia
model is a simplification from the real powertrain construction
without sacrificing the important parameters, particularly for
gear shifting study [1, 2]. All model parameters of the lumped
inertia are listed in the Appendix 1-3 including all derived
equations.
The dynamic drivetrain model consists of two major form
equations, i.e., the lumped inertia and spring-damper models.
In Fig. 4, the lumped inertia model for all mass moments of
the component inertia and the spring-damper model for all
shafts are illustrated as blocks and spring dampers,
respectively. All of these sets of equations form a chain of
drivetrain dynamic including the lock-release process of
double clutch. However, the synchronizer system equations
are described separately from Appendix 1 for better
explanation. Engine torque used in this study is simply
modeled as a static torque in a lookup table as a function of
engine speed and throttle opening percentage as shown
in Fig. 5.
The model parameters presented in Appendix 1 are mostly
based on the measurement data and the published literatures
[1, 2]. Nevertheless, in order to complete the needed
parameters in this model, some estimation values obtained
from parameter estimation tool box in MATLAB Simulink are
also used during verification of the model simulation results to
the measurement data.
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Figure 4. The 13 degree of freedom lumped inertia
drivetrain model. Blocks represent the component inertia
and springs represent the shaft stiffness. This diagram shows
a car driving with the 1st gear where the 2
nd gear is already
preselected as an example. See Appendix 1 for complete
model parameter and all the equations.
Figure 5. The static engine torque as a look up table
3.2. Synchronizer System Model
In this design, synchronizer system is used to connect two
separate rotating parts with positive interlocking made by
meshing between splines of several components as shown in
Fig. 6(a). The hub (green color) has inner and outer splines
that are meshed with the shaft and sleeve splines, respectively.
The sleeve can move in shaft axial direction from its neutral
position toward gear hub spline. The strut with a notch is
pushed by the helical spring to the outside radial direction to
plug the sleeve detent groove and hold the sleeve in neutral
position while all sleeve spline lengths are in contact only with
the hub spline. The ring has an inner cone friction surface
which rubs the outer cone friction surface of the gear hub and
acts as a synchronizer cone clutch. The ring has a narrow tab
that fills the wider slot in the hub. Thus, the ring can rotate
together with the hub and also spin relatively to it with a
limited rotating degree depending on the cone clutch torque
direction as shown in Fig. 7 (phase 2). This limited ring
rotation at its maximum locking position will block the sliding
path of the sleeve spline.
Figure 6. Synchronizer system with its (a) main
components, (b) forces involved in its splines, and (c) tip to
tip spline blocking.
The synchronizer model is derived from five phases of the
engagement processes as a function of sleeve axial
displacement from its neutral position toward the gear hub as
depicted in Fig. 7. The first, second, third, fourth, and fifth
phases are the sleeve neutral position breaking, the speed
synchronization, the ring unblocking, the hub indexing and the
spline locking, respectively.
Figure 7. Five phase of synchronizer engagement process
In the first phase, the sleeve is pushed by a hydraulic gear
lever actuator with an axial force toward the gear hub. This
initial force rises until the break through load (BTL) value is
exceeded in order to overcome the resistance of strut spring
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force that holds the sleeve on its neutral position. The BTL
can be described as [3]
(
)
where is the reaction force of strut spring, is the
friction coefficient between strut and sleeve detent groove, and
is the sleeve detent ramp angle. Before the axial force
value reaches the BTL, the sleeve brings the strut (red block in
Fig. 7) to push the ring tab and start squeezing the oil film in
cone clutch friction surface [4] and creating a cone torque
that can be described as
(
)
where is the cone mean radius, b is the width of the cone
contact surface, is the cone slip speed, h is the film
thickness, is the sleeve displacement and is the
minimum sleeve displacement for cone contact.
In the second phase, the sleeve is sliding toward the gear hub
after the axial force exceeds the BTL. This movement will
be stopped when the sleeve sliding movement is blocked by
ring spline chamfer. The tangential force in ring spline
which is a reaction of the axial force (Fig. 6(b)) creates a
chamfer torque that can be derived as
(
)
where is the chamfer pitch radius, is the friction
coefficient of the chamfer and is the chamfer angle. At the
same time, the axial force will also create a cone torque
synchronizing the targeted gear speed that can be calculated as
where and are the cone friction coefficient and angle,
respectively.
In the third phase, the ring unblocking is started when the
chamfer torque exceeds the cone torque . The axial force
through the sleeve spline chamfer rotates and pushes the
ring as a cone clutch for finishing the speed synchronization
process during ring unblocking. The cone torque at the end
of the speed synchronization is
where is the drag torque, is the freewheeling inertia
and is the angular acceleration of freewheeling
component. The drag torque is a total resistance torque
comprises all resistances from the bearings, seals, gears and
friction which can be written as
(6)
where is the gear speed, is the gear windage torque,
is the gear tooth friction torque, is the shaft drag torque,
is the clutch drag torque and is the clutch slip speed.
In the fourth phase, the sleeve continues to slide towards the
gear hub until it is blocked again by the gear hub spline
chamfer. The gear hub splines are randomly aligned to the
sleeve splines as shown in Fig. 8. Therefore, there are four
possible schemes that can be obtained in this step. The first
possibility is when the drag torque has the same direction with
chamfer torque (Fig. 8(a)). In this case, the need of axial force
to finish the spline engagement will significantly be reduced.
Meanwhile, the second possibility occurs when the drag
torque suffers the axial force to unblocking and finishes the
spline engagement (Fig. 8(b)). The phenomenon of this axial
force sudden rising is known as a second bump. Furthermore,
the third possibility shows the best alignment without any
chamfer torques (Fig. 8(c)). The worst possibility however is
achieved with tip-on-tip spline alignment. In this case, the
synchronizer will block out or fail to lock the gear hub splines
(Fig. 8(d)). The spline tip is designed as sharp as possible to
avoid a surface contact when tip-on-tip spline alignment
occurs. Thus, the sleeve still can continue to slide and unblock
the sliding way. There are four indexing torque equations
regarding to the gear hub spline alignments which can be
described in Eq. (7). The sleeve is then fully locked with the
gear hub spline in the fifth phase after passing the gear hub
splines blocking.
{
(
)
(
)
(
)
(
)
(7)
The spline chamfer angle (α) is designed to create enough
friction force on the cone friction surface to stop the slip
between the ring and the gear hub before the sleeve continues
to slide again toward the gear hub. Increasing the spline
chamfer angle will decrease the friction force on cone friction
surface and reduce the effectiveness of speed synchronization.
On the other hand, lowering the spline chamfer angle will
increase the axial force needed to unblock the synchronizer
ring blocking. In order to increase the speed synchronization
effectiveness, more friction surface areas are possible to be
created using more than one cone as used in this Getrag DCT.
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In this system, a double cone synchronizer system is utilized
for the lower gears (i.e., the 1st, 2
nd, 3
rd gears) which have a
relatively higher drag torque compared to the higher gears (the
4th
,5th
, 6th
and 7th
gears).
Figure 8. Four possibilities of the gear hub spline
indexing
3.3. Electrohydraulic System Model
The model of the electrohydraulic valve consists of three main
parts, i.e., the electromagnetic circuit, the flow equation and
the spool dynamics.
3.3.1. Electrohydraulic System
The DCT electrohydraulic system comprises several
mechanical components, e.g., mechanical oil pump, pressure
regulator valve, normally closed control valve, clutch actuator
piston and synchronizer sleeve actuator piston. Fig. 9 and Fig.
10 illustrate the DCT electrohydraulic system for clutch and
synchronizer sleeve piston actuations, respectively.
In order to actuate the clutch piston as shown in Fig. 9, electric
current is needed. This current will energize the spool in
control valve CV-C1a and open the valve. Once this valve has
been opened, the hydraulic fluid as pilot-operated pressure
will be flown through it which is generated from the oil pump
to control valve CV-C1b. The pilot-operated pressure then
pushes the valve in CV-C1b to open the path for controlling
the pressure yielded from the pressure regulator valve PRV-
C1. The pressure in clutch piston chamber is controlled only
by the pressure regulator valve. Whereas, the control valve
CV-C1a and CV-C1b are functioned as a safety valve in this
process. The simplified diagram of a clutch piston pressure
control is shown in Fig. 13. The electric current has to be
remained to control the pressure of the active clutch which is
normally open clutch type.
Figure 9. Layout of the electrohydraulic system for clutch
piston actuation
The synchronizer sleeve actuator arm can slide on two
opposite directions from its neutral position as shown in Fig.
10 for engagement process. There are four synchronizer
actuator cylinders utilized in this scheme, i.e., the first actuator
cylinder for 6th
and 4th
gears, the second actuator cylinder for
2nd
and reverse gears, the third actuator cylinder for 1st and 3
rd
gears and the fourth actuator cylinder for 5th
and 7th
gears. For
the 1st gear engagement, the sleeve actuator arm should move
toward the 1st gear. For this purpose, the electric current will
be applied to the control valve CV-G2 to open the valve and
let the hydraulic fluid flow from oil pump to gear valve GV2.
Thus, the valve in gear valve GV2 is opened. For the 1st gear
engagement, the supplied electric current will energize the
spool in the pressure regulator valve PRV-G2 for opening the
valve and controlling the pressure to the 1st-3
rd gear actuator
cylinders. After the 1st gear engagement has been finished, the
electric current will vanish from PRV-G2 and CV-G2.
Therefore, it brings the entire valves in Fig. 10 to their rest
positions and blocks all operating pressures to the actuator
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cylinders. Fig. 10 shows that the synchronizer engagement for
the next gear, i.e., gear preselect, cannot be simultaneously
performed with the synchronizer disengagement of the current
working gear, it has to be done sequentially instead.
Figure 10. Layout of the electrohydraulic system for
synchronizer sleeve piston actuation
The pressure regulator valve has one input and two output
ports as shown in Fig. 11. The valve has three main positions,
i.e., the neutral position that blocks all the flow (Fig. 11(a)),
the maximum displacement position that regulates the oil flow
from oil pump to the clutch piston (Fig. 11(b)), and the rest
position that regulates the oil flow from clutch piston to the oil
sump tank (Fig. 11(c)). The valve will move from its rest
position to its maximum displacement position
to regulate the oil flow and pressure in clutch piston chamber.
The areas of the port opening and the port closing are
a function of the valve displacement and can be described
as
[ (
)
( (
))]
[ (
⁄
)
( ( ⁄
))]
(9)
where is the port opening area as a function of ,
is the port closing area as function of , is the
radius of orifice as illustrated in Fig. 12(a) and is the
number of orifices in each port. In this case, the opening area
of orifice is not linear as shown in Fig. 12(b).
Figure 11. Pressure regulator valve with spool in (a) neutral
position (b) max. displacements and (c) min. displacements.
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The magnetic force on the spool is stated as
where is the number of coil turns, is the electric
current and ( ⁄ ) is the change in the magnetic flux
along the stroke distance with equals to ,
where is the electromagnetic inductance.
Figure 12. (a) Circular segments of charging and
discharging orifice area and (b) its nonlinear relation with
spool position.
The flow rate to the clutch piston is a nonlinear function
of the orifice area and pressure drops which can be described
as [5]
√
√
where is the discharge coefficient of the valve orifices, is
the supply pressure that is directly provided by an oil pump
depending on the engine torque and speed, is the oil sump
tank pressure, is the output pressure and is the oil density.
The spool dynamic is described by a Newton force balance
equation:
(12)
where is the spool mass, is the spool position, is
the force caused by the solenoid current , is the
viscous friction coefficient, is the spool spring constant and
is the mechanical spring force caused by the spool spring
preload. The electromagnetic valve is beneficial because of its
closed-loop control system. It leads to a loop process where
the pressure feedback is determined by an internal chamber at
each side of the spool connected by the throttles (pressure
restrictions) and of the pressure port which is then going
to the piston chamber. Thus, the proportional feedback forces
to the clutch pressure acting on the valve spool side
areas and the damping effects are provided to improve the
dynamic behavior of the spool as shown in Fig. 13.
3.3.2. Dual Clutch Dynamics
The dual clutch dynamic is highly depending on the hydraulic
pressure that pushes the clutch piston with a spring as
illustrated in Fig. 13.
3.3.2.1. Pressure Dynamics
The clutch pressure dynamic is regulated by the pressure
regulator valve and described as continuity equation of
[ ] (13)
where is the fluid bulk modulus, is the average piston
chamber volume, is the total volume flow rate, is
the volume flow rate caused by the piston movement and
equals to the change of the chamber volume , is the
volume flow rate out of the system when reaching the
maximum pressure (relief flow), is the volume flow rate
out of the system through a small discharge orifice , and
is the volume flow rate losses through the seals and
hydraulic lines. The small discharge orifice is necessary
because the spool valve does not need to cross the dead zone
to reduce the clutch pressure after the fast filling
phase, hence the oscillation and instability can be avoided.
Figure 13. Schematic diagram of the electrohydraulic wet
clutch
Page 9 of 18
3.3.2.2. Piston
The dynamic of the clutch piston is described as
where is the piston mass, is the piston friction
coefficient and is the nonlinear return spring stiffness.
There are two main types of return springs used in the modern
dual clutch transmission, i.e., inner slotted disk and multiple
round wire coil springs [6]. The multiple round wire coil
springs are used in the GETRAG 7DCI600 transmission due
to their linear behavior. , and are centrifugal forces
acting on both sides of the piston as shown in Fig. 14. The
centrifugal force for the first clutch can be described as [7]
[
] (15)
where is the rotational speed of the engine, , and are, the outer, and inner radius of the first clutch, respectively.
The return spring force for the clutch has to be greater than the
total centrifugal force acting on the piston. Typically, a
hydraulically balanced piston is used to reduce the centrifugal
force, thus lowering the required force of the return spring.
represents the unknown forces acting on the piston.
The normal force FNc1 that is converted to torque is
transmitted by the first clutch. This force is yielded by the
reaction forces between the friction plates and calculated as
(
)
where the constant parameters of , and are
identified from the measurements.
Figure 14. Schematic diagram of the dual clutch
pack assembly
3.3.3. Sleeve Actuator Dynamics
The synchronizer sleeve actuator is actuated by double action
hydraulic piston cylinders resulting in a sliding motion as
shown in Fig. 15. The hydraulic pressure in control volume
will rise to push the piston for synchronizer sleeve
actuation which can be described as [4]
∫
(
√
√ )
∫
(
√
√ )
where is the bulk modulus, is the initial hydraulic volume
with sleeve actuator at neutral position, is the discharge
coefficient, is the orifice diameter, is the solenoid
pressure, is the control volume pressure, is the
cylinder area, is the sleeve displacement, is the
cylinder diameter, is the radial clearance and is the
exhaust pressure.
Figure 15. Piston cylinder to actuate synchronizer sleeve
sliding motion
4. CONTROL OF CLUTCH AND
SYNCHRONIZER
The knowledge of the DCT system, i.e., high-quality DCT
model, is necessarily used as an observer in a control system
because some main variables cannot be measured, particularly
to control the DCT nonlinear complex system. The control
system diagram used in this study is shown in Fig. 16. This
system is integrated with an improved proportional integral
(IPI) observer to compute the system states, e.g., the unknown
forces on the clutch piston and other observed states, e.g., the
input and output from knowledge of the system [5].
The nonlinear virtual model consists of a hydraulic valve
system, dual clutches, and a synchronizer system as presented
Page 10 of 18
in this paper. The inputs of the hydraulic valve virtual model
are the input voltages for dual clutch (i.e., and ), the
input voltages for odd-even synchronizer system (i.e.,
and ), and the supply pressure . All those input
parameters are given by transmission controller unit (TCU)
which use a gradient of engine speed during synchronization
in gear shift process as control reference. Whereas, the other
inputs are provided by sensors, i.e., cylinder chamber
pressures of both dual clutches (i.e., and ), two odd
synchronizer lock-release statuses (i.e., ) and two even
synchronizer lock-release statuses (i.e., ). The outputs from
this hydraulic valve system model are the volume flow rates
for dual clutch (i.e., and ) and the volume flow rates
for odd-even synchronizer system (i.e., and ). The
inputs of the nonlinear dual clutch-synchronizer virtual model
are the same as the inputs of the hydraulic valve system that
are given by the sensor. Nevertheless, four additional inputs
from the hydraulic valve system model, i.e., , ,
and , are also utilized. The outputs from this clutch-
synchronizer virtual model are used to improve the IPI
observer robustness, i.e., the simulated clutch piston
displacements , and synchronizer sleeve
displacements , .
Figure 16. Control diagram for clutch and synchronizer
system
In order to make a seamless gear preselect, the clutch to clutch
power shift process is advanced before the sleeve synchronizer
fully locks the gear hub spline. The basic control logic of this
seamless gear preselect is similar with that of the conventional
gear preselect strategy because the synchronizer sleeve
displacement is not observed by sensor system. It attributes to
the limited capability of the sensor used in this DCT system.
The employed sensor is only sensing the final lock position of
the synchronizer sleeve. However, the advancing of the power
shift will stimulate the clutch hydraulic valve system to work
in the nearly simultaneous process with the synchronizer
hydraulic valve system due to the overlapping of work
duration time between the synchronizer hydraulic valve
system and the clutch hydraulic valve system. This proposed
process is obviously different compared to the conventional
gear preselect strategy which is using a sequential working
process. Furthermore, the backup plans used in the seamless
gear preselect control logic exhibit a different mechanism
compared to that of the conventional gear preselect strategy
control logic when the synchronizer block out occurs during
clutch to clutch power shift. In the worst condition (i.e., the
synchronizer block out), the proposed backup plans are
executed by postponing the clutch hydraulic valve action and
waiting for the synchronizer sleeve to return to its neutral
position before the sleeve slides back toward gear hub.
5. RESULT AND DISCUSSION
The seamless gear preselect simulation is achieved by
simultaneously or almost simultaneously activating the
synchronizer engagement for gear preselect with clutch to
clutch gear shifting. Due to the different characteristics of the
synchronizer and the dual clutch engagement, the process to
investigate the possibility of the seamless gear preselect is
achieved by shifting the synchronizer engagement command
signal near to the power shift command signal which is
denoted by a hydraulic fast filling action to the on-going
clutch piston chamber. Besides, the used sensors for
controlling input device are also vastly different for both
components. The pressure sensor used in the clutch
electrohydraulic system can lead the TCU to be feasible to
estimate the clutch condition during power shift. This result
differs from the sensor of the synchronizer system which can
only sense the status of sleeve final position after the
synchronizer is fully engaged. By considering these different
characteristics of the two sensors, the TCU system is only able
to fully control the dual clutch. During gear seamless
investigation by shifting the synchronizer engagement
command signal, the vehicle reactions, i.e., speed and
acceleration, are continuously observed to obtain the best
possible results, i.e., uninterrupted torque and gear shift
duration time using a genetic algorithm method. The
simulation has focused on the 1st
to 2nd
gear shifts as the most
degenerated gear shift conditions regarding to the highest
vehicle load and drag torque compared to the other upshift and
downshift cases.
During up-gear shifting, engine controller unit (ECU) reduce
the engine torque during clutch speed synchronization phase
in order to reduce the engine speed to meet lower target speed
that caused by new lower gear ratio as shown in Fig. 17.
The rotational speeds of the engine, odd clutch and even
clutches with a new seamless gear preselect are depicted in
Fig. 17. The gear shift time of the seamless gear preselect is
nearly similar with that of the sport mode gear preselect
strategy. The measurement of the shift time is started from the
beginning of the gear shift signal which is indicated by a
hydraulic fast filling action in clutch piston chamber to the
fully engagement of ongoing clutch with clutch pack.
Page 11 of 18
Figure 17. The comparison of the rotational speeds of the
engine, odd- and even clutches for two different gear
preselect strategies
Fig. 18 shows the acceleration during gear shift. For the
seamless gear preselect acceleration, the obtained acceleration
values are ranging from 10 m/s2 to 4 m/s
2 which exhibit a
similar trend with the sport mode gear preselect action. This
acceleration result is the best optimization that can be
achieved by delaying the synchronizer engagement process,
because the engine torque will be immediately interrupted
when the time delay of the synchronizer engagement process
is increased beyond that value. Furthermore, the vehicle speed
of the seamless gear preselect shown in Fig. 19 reveals
insignificant change in comparison to that of the conventional
sport mode gear preselect strategy.
Figure 18. The comparison of the vehicle accelerations for
two different gear preselect strategies
The comparison of the energy saving between these two gear
preselect strategies, i.e., the seamless and sport modes, is
obtained by simulating the vehicles in a New European
Driving Cycle (NEDC) driving scenario. All driving
parameters, e.g., driver models, are set to be equal for both
gear preselect strategies. The driving general characteristics,
i.e., gear position, braking action and throttle opening, are
quiet similar for both strategies while driving the NEDC cycle.
However, the fuel saving can still be gained by reducing the
existing wet clutch viscous drag, especially when there is a
slip in clutch after gear preselect action. The energy saving
achieved by the proposed seamless gear preselect is about
0.8% in the NEDC driving setup.
Figure 19. The comparison of the vehicle speeds for two
different gear preselect strategies
6. SUMMARY/CONCLUSIONS
The feasibility of simultaneous triggering of the clutch to
clutch power shift with a gear preselect action in order to
create a seamless method of gear preselect has been
demonstrated in this work. The structures of the DCT
mechanical and electrohydraulic systems are carefully
investigated to identify the limitation of the DCT system,
particularly for DCT Getrag 7DCI600. The advancing of
power shift trigger time toward the gear preselect is found to
be dependent on the mechanical links on the system. Thus, the
engine torque flow will be terminated when the friction
torques in both clutch and synchronizer cone clutch are
significantly decreased due to the lose contact of the friction
parts. By means of this case, the gear preselect is unattainable
to be performed simultaneously with the power shift gear
shifting in order to maintain the uninterrupted torque.
Owing an equal speed gradient of the speed synchronization
between the synchronizer and clutch system, this seamless
gear preselect is basically an enhancement of the conventional
comfort mode gear preselect strategy with the optimized
triggering power shift, i.e., performed in the condition of the
running synchronizer engagement process. However, a
simultaneous action of the synchronizer engagement and
clutch power shift can still be made by decreasing the speed
gradient of the clutch speed synchronization process, hence its
value is lower than that of the synchronizer speed
synchronization process. Furthermore, the seamless gear
preselect has shown the reduced complexity in terms of the
gear preselect logics. By using this method, the DCT system
does not have to recalculate and prepare for the next gear in its
gear preselect process because the gear preselect in this
strategy can only be performed after the gear shift signal
Page 12 of 18
exists. All these benefits of the seamless gear preselect
strategy can lead the driver to have a better driving
performance, especially for manual driving modes because the
errors of the preselected gear caused by an incorrect
estimation of the DCT system can totally be eliminated.
REFERENCES
1. Kulkarni, M., Taehyun Shim, Yi Zhang, “Shift Dynamics
and Control of Dual Clutch Transmissions”, Mechanism
and Machine Theory 42 168-182, 2007.
2. Walker, Paul D., NongZhang, “Investigation of
synchronizer engagement in dual clutch Transmission
equipped powertrains”, Journal of Sound and Vibration
331 (2012) 1398-1412, 2012.
3. Razzacki, Syed T., Jonathan E. Hottenstein,
“Synchronizer Design and Development for Dual Clutch
Transmission (DCT)”, SAE World Congress, Detroit,
Michigan, April 16-19, 2007.
4. Walker, Paul D., NongZhang, “Engagement and control
of synchronizer mechanisms in dual clutch
transmissions”, J. Mechanical Systems and Signal
Processing 26 (2012) 320-332, 2012.
5. Mustafa, R., Tobias Kassel, Gunther Alvermann, Ferit
Küçükay, “Modelling and Analysis of the
Electrohydraulic and Driveline Control of a Dual Clutch
Transmission“, Fisita 2010.
6. Mustafa, R., Ferit Küçükay, “Model-based Estimation of
Unknown Contact Forces Acting on a Piston in Dual
Clutch Transmission“, SAE Technical Paper 2012-01-
0110, 2012.
7. Alvermann, G., Ferit Küçükay, Jörgen Schulz,
“Simulation und Verifikation von
Überschneidungsschaltungen”, VDI-Schwingungen,
Dynamik und Regelung von Automatischen Getrieben,
VDI-Berichte 1917, 125-149, 2005.
8. Abdel-Halim, N. A., D. C. Barton, D. A. Crolla and A. M.
Selim, “Performance of multicone synchronizers for
manual transmissions”, Proceedings of the Institution of
Mechanical Engineers, Part D: Journal of Automobile
Engineering 2000 214: 55, 2000.
9. Bencker, Nussbaumer M., B. Schlecht, “Gearshift-
Comfort Oriented Transmission and Drive Train
Simulation at BMW”, Simpack News, Vol. 9, July 2005.
10. Bóka, G., L Lovas, J Márialigeti and B Trencséni,
“Engagement capability of face-dog clutches on heavy
duty automated mechanical transmissions with
transmission brake”, Proceedings of the Institution of
Mechanical Engineers, Part D: Journal of Automobile
Engineering 2010 224: 1125, 2010.
11. Galvagno, E., M. Velardocchia, A. Vigliani, “Dynamic
and kinematic model of a dual clutch transmission”, J.
Mechanism and Machine Theory 46 (2011) 794-805,
2011.
12. Goetz M., M. C. Levesley, D. A. Crolla, “Dynamic
Modeling of a Twin Clutch Transmission for Controller
Design”, Proceedings of the 5th
International Conference
on Modern Practice in Stress and Vibration Analysis,
Glasgow, Scotland, 9-11 September 2003.
13. Gustavsson A., “Development and Analysis of
Synchronization Process Control Algorithms in a Dual
Clutch Transmission”, Examensarbete utfört i
Fordonssystem vid Tekniska högskolan i Linköping,
2009.
14. Hackl T., Thomas Himmelbauer, Tobias Kassel, Ferit
Küçükay, “Einfachkonus-Synchronisierung mit
Servoverstärkung”, ATZ off highway, Sonderausgabe,
April 2011.
15. Haj-Fraj, A., “Dynamik und Regelung von
Automatikgetriebe”, VDI-Reihe 12, Nr. 489, VDI-Verlag,
Düseseldorf, 2002.
16. Hoshino H., “Simulation on Synchronization Mechanism
of Transmission Gearbox”, 1998 International ADAMS
User Conference, 1998.
17. Hoerbiger Antriebtecknik GmbH, “Development of
optimized friction systems for vehicle transmission”,
Hoerbiger Engineering Report 34.
18. Kim J., S Park, C Seok, H Song, D Sung, C Lim, J Kim
and H Kim, “Simulation of the shift force for a manual
transmission”, Proceedings of the Institution of
Mechanical Engineers, Part D: Journal of Automobile
Engineering 2003 217: 573, 2003.
19. Liu, Y., Datong Qin, Hoang Jiang, Charles Liu, Yi Zhang,
Zhenzhen Lei, “Shift Schedule Optimization for Dual
Clutch Transmission”, IEEE, 978-1-4244-2601-0, 2009.
20. Lovas, L., D Play, J Marialigeti, J F Rigal, “Modeling
Indexing Phase of Borg Wagner Gearbox Synchromesh
Mechanism”, International Conference Power
Transmission, 2003.
21. Lovas, L., D Play, J Marialigeti, J F Rigal, “Mechanical
behaviour simulation for synchromesh mechanism
improvements”, Proceedings of the Institution of
Mechanical Engineers, Part D: Journal of Automobile
Engineering 2006 220: 919, 2006.
22. Ni, C., Tongli Lu, Jianwu Zhang, “Gearshift control for
dry dual-clutch transmissions”, WSEAS Transactions on
Systems Issue 11, Volume 8, November 2009, 2009.
23. Rinderknecht, S., “GETRAG Power 7DCI600-Dual
Clutch Transmission for High Efficiency and Dynamics
in Inline Powertrain”, 6th
International CTI Symposium-
Innovative Automotive Transmissions, B1, 235-424
Berlin, 2007.
24. Satoh K., Masanori Shintani, Setsukazu Akai, Kazuyoshi
Hiraiwa, “Development of new synchronizer with the
lever mechanism”, JSAE 24 (2003), 93-97, 2003.
25. Schulz, J., “New Disc Spring for Dual Cluth
Transmission”, 8th
International CTI Symposium, ISBN
978-3-00-028976-7, Vol. 2, 2009.
26. Song, X., Liu Jason, Smedley Daniel, “Simulation Study
of Dual Clutch Transmission for Medium Duty Truck
Applications”, SAE Commercial Vehicle Engineering
Congress and Exhibition, Chicago, Illinois, November 1-
3, 2005.
27. Song, X., Zongxuan Sun, “Pressure-Based Clutch Control
for Automotive Transmissions Using a Sliding-Mode
Page 13 of 18
Controller”, IEEE/ASME Transactions on Mechatronics,
Vol. 17, No. 3, June 2012.
28. Walker, Paul D., Nong Zhang, Richard Tamba, “Control
of gear shifts in dual clutch transmission powertrains”, J.
Mechanical Systems and Signal Processing 25 (2011)
1923-1936, 2011.
29. Walker, Paul D., Nong Zhang, Richard Tamba, Simon
Fitzgerald, “Simulations of drag torque affecting
synchronisers in a dual clutch transmission”, Japan J.
Indust. Appl. Math. (2011) 28:119–140, 2011.
30. Xiao W., Yuxue Pan, “Virtual Design and Test of Manual
Gear Shifting Bench”, 2011 International Conference on
Electronic & Mechanical Engineering and Information
Technology, 2011 IEEE, 2011.
31. Yulong, L., Li Xingzhong, Liang Weipeng, Hanyong,
“Hydraulic System Optimization and Dynamic
Characteristic Simulation of Double Clutch
Transmission”, Procedia Environmental Sciences 10
(2011) 1065 – 1070, 2011
CONTACT INFORMATION
Mohammad Adhitya (e-mail: [email protected])
Page 15 of 18
APPENDIX
Appendix 1. Table of equations of drive train vehicle model. (all parameters symbol and constant value are further explain in
the Appendix 2 and Appendix 3)
No Equations
1
2
3 ( ) ( )
4
5 ( ) ( )
6 if both of clutches are slip
Odd side Even side
7 if odd clutch is locked
if even clutch is locked
8
{
(
) (
)
(
) [(
) (
)]
{
(
) (
)
(
) [(
) (
)]
(
) (
)
(
) (
)
Odd clutch will lock if and
Even clutch will lock if and
9
10 {
{
11
( (
)) (
) ( (
)) (
)
Above equations show the particular 1st gear and 2
nd gear in up or down gear shifting. For another gear, can be replaced by
other odd gear ratio and can be replaced by other even gear ratio.
12
13 for open synchronizer for open synchronizer
14 for open synchronizer for open synchronizer
15 for locked synchronizer for locked synchronizer
Page 16 of 18
Synchronizer lock-release process and its equations are described further in chapter 3.2 of this paper.
16 {
{
17 ((
) ) (
)
18
19
20
21
22
23 ( (
)) (
)
24
25
26
27 ( )
28
The value of constant parameter A, B and C are defined from vehicle coast-down test
Page 17 of 18
Appendix 2. Table of parameters used in drive train vehicle model.
Name of component
Dynamic parameters Constant parameters
Torque (Nm)
Angular Mass
moment of inertia
(kg m2)
Stiffness
coefficient
(Nm/rad)
Damping
coefficient
(Nms/rad)
Displa-cement
(rad)
Velocity
(rad/s)
Accele-ration
(rad/s2)
Engine
Input shaft
Flywheel
Dual mass flywheel spring
Clutch pack (flywheel side)
Friction torque of odd clutch
Odd clutch (solid shaft)
Solid shaft
Pinion gear
1st gear (or other odd gear)
Synchronizer hub
Half synchromesh
Half synchromesh
Countershaft
Final gear
Friction torque of even clutch
Even clutch (hollow shaft)
Hollow shaft
Pinion gear
2nd gear (or other even gear)
Synchronizer hub
Half synchromesh
Half synchromesh
Countershaft
Final gear
Final pinion gear
Output shaft
Propeller shaft 1st coupling
1st half of propeller shaft
Propeller shaft 2nd coupling
2nd half of propeller shaft
Differential pinion gear
Differential gear
Differential spindle
Drive shaft coupling
Drive shaft
Wheel
Page 18 of 18
Appendix 3. Table of parameters used in drive train vehicle model.
No Symbol Remarks No Symbol Remarks
1 Percentage of throttle angle 16 1st gear ratio
2 Number of odd clutch pad 17 2nd
gear ratio
3 Number of even clutch pad 18 3rd
gear ratio
4 Outer radius of odd clutch 19 4th
gear ratio
5 Outer radius of even clutch 20 5th
gear ratio
6 Inner radius of odd clutch 21 6th
gear ratio
7 Inner radius of even clutch 22 7th
gear ratio
8 A Resistance coefficient 23 Gearbox final gear
9 B Resistance coefficient 24 Differential gear ratio
10 C Resistance coefficient 25 Vehicle weight
11 Odd clutch normal force 26 Vehicle velocity (m/s)
12 Even clutch normal force 27 Road angle
13 Kinetic friction coefficient 28 Vehicle velocity (km/h)
14 Static friction coefficient 29 Vehicle drag coefficient
15 Rolling coefficient 30 Wheel radius
